1. Field of the Invention
The present invention relates to a planetary gear type transmission mechanism, for example, for an automatic transmission, which can provide greater freedom in selection of speed ratios in designing a multi-speed-ratio automatic transmission.
2. Description of the Related Art
Japanese Patent Application Laid Open No. 52-149562 discloses a planetary gear type transmission mechanism. FIG. 1 shows the construction of the conventional planetary gear type transmission mechanism disclosed in the above-mentioned publication. As shown in FIG. 1 the disclosed planetary gear type transmission mechanism includes first, second and third planetary gear sets G1, G2 and G3 interposed coaxially between an input shaft I and an output shaft O. Each of the first planetary gear set G1 near the input shaft I, the third planetary gear set G3 near the output shaft O, and the second planetary gear set G2 therebetween is a single pinion type planetary gear set, which comprises a sun gear S1, S2 and S3, an internal ring gear R1, R2 and R3, a planet pinion P1, P2 and P3 meshing with the corresponding internal ring gear and the corresponding sun gear, and a pinion carrier PC1, PC2 and PC3 rotatably supporting the corresponding planet pinions. The sun gears S2 and S3 of the second and third planetary gear sets G2 and G3 are integrally coupled with each other to form a first rotating member 1. The first rotating member is connectable with the input shaft I through a first clutch C1. The pinion carrier PC3 of the third planetary gear set G3 is integrally coupled with the output shaft O to form a second rotating member 2. The pinion carrier PC2 of the second planetary gear set G2 and the internal ring gear R3 of the third planetary gear set G3 are integrally coupled with each other to form a third rotating member 3. The third rotating member 3 is able to be fixed to a transmission casing 7 by means of a first brake B1 and is also connectable with the input shaft I through a second clutch C2. The pinion carrier PC1 of the first planetary gear set G1 and the internal ring gear R2 of the second planetary gear set G2 are integrally coupled with each other to form a fourth rotating member 4 which is adapted to be fixed on the transmission casing 7 by means of a second brake B2. The sun gear S1 of the first planetary gear set G1 is integrally coupled with, the input shaft I to form a fifth rotating member 5. The internal ring gear R1 of the first planetary gear set G1 forms a sixth rotating member 6 and is able to be fixed on the transmission casing as a seventh rotating member 7 by means of a third brake B3.
The alignment chart showing relationships among rotation speed ratios of the rotating members in the foregoing construction of the transmission mechanism is illustrated in FIG. 2. In FIG. 2, the positions 1 to 4 on the horizontal axis show relative positions of the first to fourth rotating members determined corresponding to the set gear ratios (ratio of gear teeth number of the sun gear versus gear teeth number of the internal ring gear) and positions 5, 4, 6 show relative position of the fifth, fourth and sixth rotating member determined corresponding to the set gear ratios. Vertical axes extending across the points 1 to 4 and 5, 4, 6 represent rotation speed ratio of the relevant rotating members (ratio of rotation speed of the rotating member versus the rotation speed of the input shaft). Both alignment charts of the first to fourth rotating members 1 to 4 and of the fifth, fourth and sixth rotating members 5, 4, 6, are illustrated on the same figure. The rotation speed ratio 0 represents a state in which the relevant rotating member is fixed, 1 represents the state in which the relevant rotating member rotates in the same direction as the input shaft I (forward direction) at the same speed as that of the input shaft, and -1 represents the state in which the relevant rotating member rotates in the opposite direction to the input shaft (reverse direction) at the same speed as that of the input shaft. As shown in FIG. 2, the disclosed planetary gear type transmission mechanism establishes six forward speed ratios and a single reverse speed ratio by engaging two friction elements of the clutches C1, C2 and the brakes B1, B2, B3 to limit rotation of the corresponding rotating members. If a rate of the distances between 1 and 2, 2 and 3, and 3 and 4 is expressed as 1:A:B, and a rate of distances between 5 and 4, and 4 and 6 is expressed as 1:C, the gear ratios .alpha..sub.1, .alpha..sub.2 and .alpha..sub.3 (ratio of gear teeth number of the sun gear versus gear teeth number of the internal ring gear) of the first, second and third planetary gear sets G1, G2 and G3 can be expressed respectively by: EQU .alpha..sub.1 =C (1) EQU .alpha..sub.2 =B.div.(1+A) (2) EQU .alpha..sub.3 =A (3)
The following Table 1 represents the relationships between the friction elements to be engaged (refer to ".largecircle." marks) and the gear positions established thereby as well as the speed ratios at respective gear positions.
TABLE 1 ______________________________________ Gear Friction element position C1 C2 B1 B2 B3 Speed ratio ______________________________________ Forward 1st .largecircle. .largecircle. ##STR1## 2nd .largecircle. .largecircle. ##STR2## 3rd .largecircle. .largecircle. ##STR3## 4th .largecircle. .largecircle. 1 5th .largecircle. .largecircle. ##STR4## 6th .largecircle. .largecircle. ##STR5## Reverse .largecircle. .largecircle. ##STR6## ______________________________________
In the prior art set forth above, respective speed ratios other than the speed ratio at the fourth gear position (speed ratio=1) are determined by the three parameters A, B and are C and correlated to each other. This makes it impossible to determine the speed ratio at each gear position independently of other gear positions. Therefore, although desired speed ratios can be realized as one likes in the alignment chart, by determining the parameters A, B and C to make speed ratios near the desired values, the transmission mechanism of the desired speed ratios is difficult to realize according to desired speed ratios in the practical implementation, since it is known to those skilled in the art that unless the gear ratios .alpha..sub.1, .alpha..sub.2 and .alpha..sub.3 of the planetary gear sets G1, G2 and G3 determined by the parameters A, B and C are maintained within a predetermined range, the transmission mechanism cannot be applied for practical use. In practice, the preferred range of gear ratios of a planetary gear set is 0.35 to 0.6. If the gear ratio exceeds 0.6, due to an upper limitation of the size of the automatic transmission casing the diameter of the planet pinion becomes unacceptable small, which requires that the planet pinion must be rotated at a high speed, and the size of the gear teeth thereof become unacceptably small and potentially lower durability of the planetary gear set to an unacceptable level for practical use. On the other hand, when the gear ratio becomes less than 0.35, because of restrictions on the size of the transmission casing, the diameter of the sun gear tends to become excessively small and may cause lack of strength of a shaft extending through the sun gear.
In the practical example, when the automatic transmission is designed with parameters A=0.420, B=0.682 and C=0.887, to establish the following speed ratios: first gear position=3.38, second gear position=1.91, third gear position=1.34, fourth gear position=1.00, fifth gear position=0.75, sixth gear position=0.62, and reverse gear position=-3.45, the gear ratios calculated from said expressions (1) to (3) become .alpha..sub.1 =0.887, .alpha..sub.2 =0.48 and .alpha..sub.3 =0.42. In this example, the gear ratio .alpha..sub.1 exceeds the above-mentioned preferred range. Therefore, the speed ratio sought in this example cannot be realized.
Another type of a planetary gear type transmission mechanism is also well-known, as shown in FIG. 3. This prior art differs from the prior art in FIG. 1 in the following points. The third brake B3 in FIG. 1 as the friction element for fixing and releasing the internal ring gear R1 as the sixth rotating member to and from the transmission casing as the seventh rotating member is deleted, and in FIG. 3 the internal ring gear R1 is integrally coupled with the transmission casing as the seventh rotating member 7. In FIG. 3, moreover, the input shaft I forms the fifth rotating member 5 by itself and the sun gear S1 forms the sixth rotating member 6 by itself, between which rotating members 5 and 6 is provided a third clutch C3 as a friction element for connecting and disconnecting these rotating members 5 and 6.
In this arrangement in FIG. 3, as shown in FIG. 4 the alignment chart showing relationships among rotation speed ratios of the rotating members is same as that shown in FIG. 2 of the prior art in FIG. 1. Therefore, as shown in the following Table 2 representing the relationship between the friction elements to be engaged (refer to ".largecircle." marks) and the gear positions established thereby, the prior art in FIG. 3 also establishes speed ratios at each gear position similar to those of the first prior art in FIG. 1.
TABLE 2 ______________________________________ Gear Friction element position C1 C2 C3 B1 B2 Speed ratio ______________________________________ Forward 1st .largecircle. .largecircle. ##STR7## 2nd .largecircle. .largecircle. ##STR8## 3rd .largecircle. .largecircle. ##STR9## 4th .largecircle. .largecircle. 1 5th .largecircle. .largecircle. ##STR10## 6th .largecircle. .largecircle. ##STR11## Reverse .largecircle. .largecircle. ##STR12## ______________________________________
In the prior art in FIG. 3, if a rate of distances between 1 and 2, 2 and 3, and 3 and 4 is expressed as 1:A:B, and a rate of distances between 6 and 4, and 4 and 7 is expressed as 1:C, the gear ratios .alpha..sub.1, .alpha..sub.2 and .alpha..sub.3 (ratio of gear teeth number of the sun gear versus gear teeth number of the internal ring gear) of the first, second and third planetary gear sets G1, G2 and G3 can be expressed respectively by: EQU .alpha..sub.1 =C (1) EQU .alpha..sub.2 =B.div.(1+A) (2) EQU .alpha..sub.3 =A (3)
and these gear ratios .alpha..sub.1, .alpha..sub.2 and .alpha..sub.3 are similar to the prior art in FIG. 3.
When the values given for the example of FIG. 1 are also applied to the prior art in FIG. 3, the gear ratios .alpha..sub.1, .alpha..sub.2 and .alpha..sub.3 respectively become 0.887, 0.48 and 0.42 Therefore, similarly to the example of FIG. 1, the gear ratio .alpha..sub.1 of the first planetary gear set G1 goes out of the preferred range (0.35 to 0.6) and the same drawback as that in the example of FIG. 1 is caused.